JP3186452B2 - Suspension control device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a method of controlling a damping coefficient of a variable damping force shock absorber based on at least the amount of movement caused by a change in posture of a vehicle body, the speed thereof, or the magnitude of movement input such as acceleration thereof. And a suspension control device.

[0002]

2. Description of the Related Art A conventional semi-active suspension control device is disclosed in, for example, Japanese Patent Application Laid-Open No. 3-42319.
Is described in Japanese Unexamined Patent Application Publication No. 2000-205,878. In this conventional example, at least a small damping force in the extending direction (hereinafter, simply referred to as “extension side”) and the compression side damping force in the compression direction (hereinafter, simply referred to as “compression side”) are respectively reduced by input of a control signal. Shock absorber capable of changing force (hereinafter simply referred to as low damping force) and large damping force (hereinafter simply referred to as high damping force), and sprung speed measuring means for measuring sprung speed corresponding to the vehicle body And relative speed measuring means for measuring a relative speed between the sprung portion and the unsprung portion corresponding to the wheel side,
Sign determining means for determining whether the sign of the sprung speed and the sign of the relative speed match or not, and when both signs match and the sign of the relative speed is positive, the extension side has a high damping force and the compression side has a high damping force. When the two signals agree with each other and the sign of the relative speed is negative, a control signal is output to decrease the damping force on the extension side and increase the damping force on the compression side. When there is a disagreement,
Control signal output means for outputting a control signal for reducing the damping force on both the extension side and the compression side. Since the damping force is specifically expressed as a product of a piston speed and a damping coefficient of the piston installed in the shock absorber, it is strictly set for the shock absorber that:
It is reasonable to assume that the damping coefficient excludes the piston speed which is a parameter.

However, in this conventional example, the high damping coefficient and the low damping coefficient set on the extension side and the compression side in each damping force variable shock absorber can be set only to constant values. That is, in each damping force variable shock absorber used in this suspension control device, the high damping coefficient specifically set on the extension side and the compression side is a constant value, and when the extension side is set to this fixed high damping coefficient, the compression side Is set to a constant low damping coefficient, and when the compression side is set to a constant high damping coefficient, the extension side is set to a constant low damping coefficient. . That is, in this damping force variable shock absorber, the respective damping coefficients on the extension side and the compression side can be set to only three positions.

On the other hand, the so-called skyhook theory has attracted attention from the viewpoint of the vibration control effect and attitude control of the vehicle body. In order to achieve the Skyhook theory in a vehicle in accordance with the so-called Karnopp's rule, the amount of behavior occurring in the vehicle, specifically, the amount of vehicle It must be possible to continuously change and set the damping force of each shock absorber for the vehicle motion state quantity such as the speed or the acceleration. Therefore, the present applicant has previously proposed a suspension control device using a variable damping force shock absorber described in Japanese Patent Application No. 5-328426. Briefly describing the variable damping force shock absorber used in these suspension control devices, a disc valve, a reed valve, and the like are provided between a piston provided in each shock absorber and a valve element provided in each piston. When the expansion-side fluid path and the compression-side fluid path that are automatically opened and closed are formed, and the valve body is rotated or moved relative to the piston by the actuator, the expansion-side fluid path and the compression-side fluid path intervene as orifices. Since the opening area between the piston and the valve body of each fluid path is changed, the control amount to this actuator is changed and controlled to restrict the variable orifice (flow resistance, and at the same time this damping force variable (Corresponding to the damping coefficient of the damping force changed and controlled by the shock absorber) Damping force (ie where is the attenuation coefficient) can be continuously changed individually controlled.

When the damping coefficient on the extension side is set to a relatively high damping coefficient, the damping coefficient on the compression side becomes a low damping coefficient.
When the compression-side damping coefficient is set to a relatively high damping coefficient, the extension-side damping coefficient becomes a low damping coefficient itself, which is the same or almost the same as that of the conventional example. Alternatively, the pressure-side damping coefficient can be continuously increased or decreased. Further, a step motor is specifically used as the actuator, and the rotation amount of the step motor, that is, the number of steps (more strictly, the number of pulses of the control signal) is used as the control amount. That is, at least the damping force (= damping coefficient) on the high damping side is
It has a unique relationship with the relative rotation angle of the valve element that is linearly related to the rotation angle of the step motor, that is, the rotation position.

In such a suspension control apparatus using a shock absorber having a continuously variable damping force, in order to simply realize the Karnopp's law, for example, the vehicle body side spring up / down The speed is calculated or detected. Specifically, as the sprung vertical speed increases in the positive region, the extension-side damping force is gradually increased, and as the sprung vertical speed decreases in the negative region, the compression-side damping force is gradually increased. I have to. When tuning to a specific vehicle, in order to achieve a smooth ride, especially in a low-speed running state where the vehicle speed is low, the region where the absolute value of the sprung vertical speed is small, that is, the displacement of the vehicle body occurs slowly. In the range where the absolute value of the sprung vertical velocity is less than or equal to this low damping threshold, the damping force of the variable damping force shock absorber is set to be as low as possible on both the extension side and the compression side. are doing.

[0007]

The above-mentioned body movement state resulting from the change in the posture of the vehicle body includes a roll movement accompanying a turn or the like. This roll motion is generally caused by a roll moment resulting from a difference in an action point between a cornering force of a wheel at the time of turning or the like and a lateral acceleration acting on the vehicle body. The absorber is on the compression side, and the shock absorber for the turning inner ring is on the extension side. Then, in order to prevent the change in the posture of the vehicle body against such a roll motion,
Since the damping force by the variable damping force shock absorber is given by the product of the piston speed and the damping coefficient as described above, the damping coefficient on the extension side and the compression side of the damping force variable shock absorber is determined by the direction and magnitude of the roll speed. It is proposed that the setting be made according to the above. More specifically, an attempt is made to set the damping coefficient of the variable damping force shock absorber so that the damping coefficient on the extension side or the compression side in the matched direction uniquely increases with an increase in the roll speed. Here, in order to facilitate understanding, it is assumed that the damping coefficient on the extension side or the compression side is simply set to increase linearly as the roll speed increases.

However, since the setting of the damping coefficient of the variable damping force shock absorber by the roll suppression control depends on the magnitude of the roll speed, the damping force generated by the variable damping force shock absorber by the roll suppression control. Is the product of the roll speed and the damping coefficient obtained by multiplying the roll speed by the scalar, and consequently depends on the square value of the roll speed. On the other hand, the phase of the damping force shifts or the amplitude changes. Therefore, the roll suppression control is likely to give an uncomfortable feeling to at least a driver who is accustomed to the conventional shock absorber.

In order to solve such a problem, the damping coefficient of the variable damping force shock absorber may be set to a certain value which is somewhat higher on both the extension side and the compression side, similarly to that of the conventional shock absorber. However, in the variable damping force shock absorber having the above-described configuration, the actual situation is that the damping coefficients on the extension side and the compression side cannot be set to a certain high value at the same time.

[0010] The present invention has been developed in view of these problems, and provides a suspension control device that enables roll suppression control without giving a driver who is accustomed to conventional shock absorbers a sense of incongruity. The purpose is to.

[0011]

As shown in FIG. 1, a suspension control device according to the present invention is interposed between a vehicle body-side member and a wheel-side member, and responds to input control signals. By controlling the position of the valve body by an actuator driven in accordance with the variable damping force, a damping force variable shock absorber capable of setting the damping coefficient on either the extension side or the compression side to be large or both can be set small, and the posture of the vehicle body Calculating and setting a damping coefficient for suppressing a change in the posture of the vehicle body based on at least a vehicle body movement state detection value detected by the vehicle body movement state detection means; The control signal is output to the actuator such that the actual position of the valve body matches the target position of the valve body corresponding to the damping force coefficient, and the damping force variable shock absorber is output. A control unit for controlling a damping coefficient of a sober, wherein the vehicle body movement state detection unit includes at least a roll movement state detection unit that detects a roll movement state occurring in the vehicle, and the control unit includes: in order to suppress at least the rolling motion state based on the roll movement state detection value detected by the roll motion state detecting means, is set the damping coefficient of the damping force control shock absorber comprising a pressure side in advance
Which comprising the rolling motion control means for setting a constant predetermined extension side damping coefficient is preset damping coefficient of the damping force control shock absorber comprising a extension side and sets a constant predetermined pressure side damping coefficient It is.

[0012]

According to the suspension control device of the present invention, as shown in the basic configuration diagram of FIG. 1, the roll motion state detecting means provided as the vehicle body motion state detecting means for detecting the vehicle body motion state caused by the change in the posture of the vehicle body is provided with: A roll motion state generated in the vehicle is detected, and the roll motion control means provided in the control means determines at least the roll motion state based on the roll motion state detection value detected by the roll motion state detection means. in order to suppress the body attitude change settings, for example, be in a roll state of motion of the cornering, the pressure side damping coefficient constant where the damping coefficient is preset turning outer wheel side damping force variable shock absorber comprising a pressure side was set at a constant predetermined extension side damping coefficient of the damping coefficient is preset turning inner side of the damping force control shock absorber serving as expansion side as well as, those The control means controls at least the actual position of the valve body to coincide with the target position of the valve body corresponding to the damping coefficient set as described above in order to suppress at least a change in the vehicle body posture due to the roll motion state. Since the damping coefficient of each damping force variable shock absorber is controlled by outputting a signal to the actuator, for example, in a roll motion state at the time of turning, the damping force generated by the damping force variable shock absorber of the turning outer wheel is: Since the damping force which appears as a product value of the predetermined compression-side damping coefficient and the roll speed and which is expressed by the variable damping force shock absorber of the turning inner wheel appears as a product value of the predetermined extension-side damping coefficient and the roll speed, each of the predetermined values is obtained. If the damping coefficient is set to the same value as that of a conventional shock absorber, it becomes easier to get used to a vehicle equipped with the conventional shock absorber. To the driver without giving any special discomfort, it is set larger than that of the predetermined damping coefficient conventional shock absorbers are set in the variable damping force shock absorber, suppressing the effect of preventing roll movement is improved.

[0013]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the suspension control device of the present invention will be described below with reference to the drawings. In the suspension control device of the present embodiment, the bounce suppression control and the roll suppression control are executed in parallel as the vehicle body posture change suppression control. FIG. 2 is a schematic configuration diagram showing an embodiment of the present invention, and includes a variable damping force shock absorber 3FL constituting a suspension device between each of the wheels 1FL-1RR and the vehicle body 2.
Step motor 41FL for switching the damping force of these variable damping force shock absorbers 3FL-3RR.
To 41RR are controlled by a control signal from the controller 4 described later.

Each of these variable damping force shock absorbers 3FL-3RR is a twin-tube gas-filled strut type having a cylinder tube 7 composed of an outer cylinder 5 and an inner cylinder 6, as shown in FIGS. And the upper and lower pressure chambers 9U, 9
L. 4 to FIG.
As is particularly apparent in FIG. 7, a seal member 9 which is in sliding contact with the inner cylinder 6 is molded on the outer peripheral surface and a center opening 10 is formed on the inner peripheral surface.
, And an upper half 12 fitted inside the lower half 11.

The lower half body 11 is provided with an extension oil flow path 13 penetrating vertically, and the lower half body 11 is provided extending from the upper surface side downward to the lower side of the seal member 9. A hole 14a having a larger diameter than the extension-side oil flow path 13 and a hole 14b drilled from the outer peripheral surface of the cylindrical body 11 to the bottom of the hole 14a.
, The annular grooves 15U, 15L formed at the upper and lower open ends of the center opening 10, the annular groove 15U formed on the upper surface side, and the expansion-side oil flow path 13. A long groove 16 communicating with each other, and an annular groove 15L formed on the lower surface side
A long groove 17 is formed to communicate with the lower oil passage 13, the lower end of the expansion oil passage 13 and the long groove 17 are closed by the expansion disk valve 18, and the upper end of the compression oil passage 14 is closed by the compression disk valve 19.

The upper half 12 has a small-diameter shaft portion 21 inserted into the center opening 10 of the lower half body 11 and the shaft portion 2.
The lower end face of the small-diameter shaft portion 21 is formed at the center of the small-diameter shaft portion 21 and the large-diameter shaft portion 22. 23a extending from the side to the middle part of the large-diameter shaft portion 22, and a hole 23b communicating with the upper end side of the hole 23a and having a smaller diameter than this.
And a hole 23c having a larger diameter than the hole 23c communicating with the upper end side of the hole 23b is formed. Pair of through holes 24 penetrating in the direction toward the inner peripheral surface side
a, 24b and 25a, 25b are drilled, and the upper end side of the hole 23a of the large-diameter shaft portion 22 is connected to the arc-shaped groove 2 communicating therewith.
6 is formed, and an L-shaped pressure-side oil flow path 27 that communicates the arc-shaped groove 26 with the lower end face is formed.
Is blocked by

The lower half 11 and the upper half 12 are positioned below the lower half 11 of the small-diameter shaft 21 with the small-diameter shaft 21 inserted into the central opening 10 of the lower half 11. The nut 29 is screwed to the lower end protruding from the nut, and the nut 29 is tightened to be integrally connected. Further, a cylindrical valve element 31 having a closed upper end, which constitutes a variable throttle, is rotatably disposed in the hole 23a of the upper half body 12. As shown in FIG. 4, the valve body 31 has an upper half 1
2, a through-hole 32 is formed at a position facing the arc-shaped groove 26 of the large-diameter shaft portion 22 so as to reach the inner peripheral surface in the radial direction, and as shown in FIGS. A communication groove 33 is formed in the outer peripheral surface corresponding to the portion between the through holes 24a and 24b of the portion 21 to communicate them with each other. Further, as shown in FIG.
And 25b, an elongated hole 34 extending in the axial direction is formed on the outer peripheral surface corresponding to the inner peripheral surface side.
The positional relationship between the through hole 32, the communication groove 33, and the elongated hole 34 is determined by the damping force characteristic (proportional to the damping coefficient) with respect to the rotation angle of the valve body 31 shown in FIG. Is selected to obtain

That is, for example, at the position A in FIG. 8, which is the maximum rotation angle position in the clockwise direction, as shown in FIG. For the pressure side movement, lower pressure chamber 9L
, A pressure side flow path C1 (shown by a dashed line) passing through the orifice formed by the opening end and the pressure side disk valve 19 to the upper pressure chamber 9U, through the pressure side oil flow path 14, and the lower pressure chamber 9L.
Through the inner peripheral surface of the valve body 31, the through hole 32, the arc-shaped groove 2
6. A pressure-side flow path C2 (shown by a broken line) that passes through the pressure-side oil flow path 27, passes through an orifice formed by the opening end thereof, and the pressure-side disk valve 28, and goes to the upper pressure chamber 9U, is formed. For the extension side movement, the upper pressure chamber 9U passes through the elongated groove 16 and the extension side flow path 13 and passes through an orifice formed by the opening end and the extension side disc valve 18 to the broken line toward the lower pressure chamber 9L. Only the expansion side flow path T1 shown in the figure is formed, and a high damping force is generated on the expansion side, which rapidly increases in accordance with an increase in the piston speed, and a low damping force on the compression side is slightly increased in accordance with the increase in the piston speed. Generates damping force.

By rotating the valve element 31 in the counterclockwise direction from the position A, the communication groove 33 of the valve element 31 and the through holes 24a, 25a of the small diameter shaft section 21 are in communication with each other, as shown in FIG. The opening area between the communication groove 33 and the through holes 24a and 25a gradually increases with an increase in the rotation angle. For this reason, as shown in FIG. 5A, the elongated groove 16, the annular groove 15 </ b> U,
Through hole 24a, communication groove 33, through hole 24b, annular groove 1
5L, a flow path T2 that passes through the long groove 17 and passes through the orifice formed by the long groove 17 and the pressure-side disc valve 18 toward the lower pressure chamber 9L is formed, and the maximum value of the damping force is shown in FIG. In this way, as shown in FIG. 5B, the flow path C1 and the flow path C1 are gradually reduced as the opening area between the communication groove 33 and the through-holes 24a and 25a of the small-diameter shaft portion 21 increases. In order to maintain the state where C2 is formed,
Maintain the minimum damping force condition.

Further, when the valve body 31 is rotated counterclockwise to be in the vicinity of the position B, as shown in FIG. . For this reason, as shown in FIG. 6A, the piston 8 passes through the long groove 16, the annular groove 15U, the through hole 25a, the long hole 34, and the hole 23a in parallel with the flow paths T1 and T2 as shown in FIG. As a result, a flow path T3 toward the lower pressure chamber 9L is formed, the extension-side damping force becomes the minimum damping force state, and a hole in addition to the flow paths C1 and C2 is provided for the piston 8 on the compression side. Part 23a, long hole 34, through hole 25a,
The flow path C3 that reaches the long groove 16 through the annular groove 15U, the hole 23a, the long hole 34, the through hole 25b, and the annular groove 15
L, a flow path C4 that reaches the long groove 16 through the through hole 24b, the communication groove 33, the through hole 24a, and the annular groove 15U is formed, but maintains the minimum damping force state as shown in FIG.

Further, when the valve element 31 is rotated in the counterclockwise direction, the opening area between the elongated hole 34 and the through holes 24b and 25b is reduced, and the elongated hole 34 and the through hole 24 are rotated at a rotation angle θ _B2.
Although the gap between b and 25b is cut off as shown in FIG. 7, the opening area between the through hole 32 and the arc-shaped groove 26 gradually decreases from the rotation angle θ _B2 . For this reason, during the period from the rotation angle θ _B2 to the maximum rotation angle θ _{C in the} counterclockwise direction, the minimum damping force state is maintained because the flow paths T1 and T2 coexist for the movement of the piston 8 on the extension side. Conversely, when the piston 8 moves on the pressure side, the opening area between the through hole 32 and the arc-shaped groove 26 gradually decreases, so that the maximum damping force gradually increases, and the valve body 31 moves to the position C. When the piston arrives, as shown in FIG. 7, the space between the through hole 32 and the arc-shaped groove 26 is cut off, so that the lower pressure chamber 9 is moved with respect to the pressure side movement of the piston.
The flow path from L to the upper pressure chamber 9U is only the flow path C1, and the pressure side is in a high damping force state.

That is, since these damping force characteristics are determined by the opening areas of the orifices formed between the valve element 31 and the piston 8, the valve element 31 is rotated relative to the piston 8. The rotation angle of the step motor to be controlled is a control amount for selectively setting the flow resistance determined by the throttle of the orifice, that is, the damping coefficient.
Each of the damping forces is expressed in the form of a product obtained by multiplying the damping coefficient by the piston speed.

Therefore, the rotation angle of this step motor is
Assuming the position P, the damping force on the extension side becomes the maximum
Position P is the maximum position P on the extension side_TMAXAnd pressure side
The position P at which the damping force of the
Condition P_CMAXHowever, here, for convenience,
Intermediate value in the range where both damping force and compression damping force are set to low damping force
Is set to "0", and the extension side damping force is high.
Positive position change in the direction to become positive and damping force on the compression side
If the position change in the direction where
Extension side maximum position P_TMAXIs a positive sign and simply P_MAXAnd expressed
Pressure side maximum position P_CMAXIs a minus sign and simply (-P
_MAX). However, the absolute position of each of these maximum positions
Logarithmic value | P_MAX| Does not necessarily have to be the same value. So
Then, the negative pressure side maximum position (-P_MAx)
The maximum position P on the extension side that becomes a positive value from_MAXTotal attenuation up to
Positive threshold for position P sandwiching "0" in the force control range
Value P _T1From the negative threshold P_C1Up to the extension side low damping force D
/ F_T0And compression side low damping force D / F_C0The performance
Soft range to achieve smoothness especially in low speed driving
(Hereinafter simply referred to as SS range)
The range where the position P is large in the positive direction, that is, the position P
Is the positive threshold P_T1From the positive maximum extension position P
_MAxUp to the extension side control where the extension side damping force is set high.
(Hereinafter simply referred to as HS range).
The range where the position P is smaller in the negative direction, that is, the position
P is the negative threshold P_C1Negative pressure side maximum position from
(-P_MAx), The compression side damping force is set high.
Control range (hereinafter, also simply referred to as SH range).
You. Therefore, the positive threshold P_T1With the positive low attenuation threshold
And the negative threshold P_C1As a negative low attenuation threshold.

Although the details will be described later, FIG.
The first predetermined extension side damping force D / F _T2 and the second predetermined extension side damping force D / F _T3, the first and second predetermined extension side damping coefficient of each C _T2, C _T3 are usually of the front and rear wheels as shown in Is set so as to be equal to or substantially equal to the extension side damping coefficient which is a constant value of the conventional shock absorber used for the first and second positions, and the positions P which achieve the predetermined extension side damping coefficients C _T2 and C _T3 are the first position and the position P, respectively. corresponds to a second predetermined extension side position value P _T2, P _T3. The first predetermined pressure side damping force D / F _C2 and the second predetermined pressure side damping force D / F _C3 shown in FIG. 8, the respective first and second
Predetermined pressure side damping coefficient C _C2, C _C3 are and have set to be equal to or substantially equal to the compression side damping coefficient is a constant value of conventional shock absorber for use in conventional front and rear wheels, each place pressure side damping coefficient C _C3 , C _{C3 correspond} to the first and second predetermined pressure side position values P _C3 , P _C3 , respectively. Further, as described in detail below, when a roll motion occurs, the front-rear roll axis passing through the center of gravity of the vehicle (which has a different meaning from a roll axis connecting a so-called roll center) is not changed vertically and horizontally. the first predetermined extension side damping coefficient C _T2 and the first predetermined pressure side damping coefficient C _C2, are set such that the absolute value to each other is equal to or almost equal, and the second predetermined extension side damping coefficient C _T3 Second
The predetermined pressure side damping coefficients C _C3 are also set so that their absolute values are equal or almost equal to each other. Further, according to the damping force characteristic (damping coefficient characteristic) of FIG. 8, at each of the predetermined position values for achieving the predetermined expansion-side damping coefficient and the predetermined compression-side damping having the same absolute value, the absolute value of the predetermined expansion-side position value is obtained. The value is slightly smaller than the absolute value of the predetermined pressure side position value.

On the other hand, a cylindrical piston rod 35 is fitted into the hole 23c of the upper half body 12, and the upper end of the piston rod 35 projects upward from the cylinder tube 7, as shown in FIG. The upper end side is the vehicle body side member 36.
Is fixed to a bracket 37 attached to the bracket 37 via nuts 39 via rubber bushes 38U and 38L.
The rotating shaft 41a and the above-described valve body 31 are connected by a connecting rod 42 loosely inserted into the piston rod 35. 43 is a bumper rubber. The lower end of the cylinder tube 7 is connected to a wheel-side member (not shown).

The controller 4 has, on its input side,
As shown in the figure, according to the vertical acceleration provided on the vehicle body side corresponding to each wheel position, the vertical acceleration detection values X _2FL ″ -X composed of analog voltages that are positive in the upward direction and negative in the downward direction
Vertical acceleration sensors 51FL to 51RR as vertical acceleration detecting means for outputting _2RR "are connected, and the damping force of each of the damping force variable shock absorbers 3FL to 3RR is a damping force on the output side (the control amount of the output terminal is a damping coefficient). Are connected.

The controller 4 includes a microcomputer 56 having at least an input interface circuit 56a, an output interface circuit 56b, an arithmetic processing device 56c and a storage device 56d;
An A / D converter 57FL~57RR supplied to the input interface circuit 56a FL～51RR the vertical acceleration detection value X _2FL "~X _2RR" is converted into a digital value, each step motor which is output from the output interface circuit 56b 41
A step control signal for FL to 41RR is input, and is converted into a step pulse to convert each step motor 41FL to 41RR.
The motor drive circuits 59FL to 59RR for driving the 41RR are provided.

Here, the operation of the microcomputer 56
The processing device 56c performs each of the above operations by an arithmetic process described later.
Lower acceleration detection value X_2FL"~ X_2RRIntegrate ″
Speed (also referred to as sprung vertical speed) X_2FL'~ X_2RR'
Calculated and each sprung vertical speed X described later_2FL'~ X_2RR'
The dead zone threshold value (−X_2i0') ~ X
_2i0'Each sprung point except within the dead zone between (i = FL to RR)
Lower speed X_2FL'~ X_2RR', Each sprung vertical speed
X_2FL'~ X_2RR'' For bounce motion suppression control according to
Damping force D / F to each damping force variable shock absorber
To achieve 3FL to 3RR, the damping force variable shock
Each step motor 41FL-41 of absorber 3FL-3RR
The target rotation angle that determines the RR damping coefficient, that is, the target position of the valve
(Hereinafter simply referred to as bounce control target position)
Write) P_DBi(I = FL ~ RR)
The sprung vertical speed X of the left and right wheels_2FL'~ X_2RR'Deviation
To the front wheel side and rear wheel side roll speed R_RF, R_RRCalculate
And each roll speed R_RF, R _RRDepending on the direction and size of
Decrease the damping force D / F for the roll motion suppression control
To achieve with variable damping shock absorbers 3FL-3RR
Each of the damping force variable shock absorbers 3FL to 3RR
An eye for determining the damping coefficient of the step motors 41FL to 41RR
The target rotation angle, that is, the target position of the valve body (hereinafter simply referred to as low
Control target position) P_DRi(I = FL ~ RR)
And set both target positions P_DBi, P_DRiSuitable for
Weighting (distribution ratio) and the final
Target position P_DCalculate and set the target position
P_DAnd current position P_ACalculate the difference between
Motor control circuits 59FL-59
RR, the rotation angle of the step motor,
Variable damping force variable shock absorber 3 according to position
Open loop control of FL-3RR damping force (= damping coefficient)
I do.

The storage device 56d stores a program necessary for the arithmetic processing of the arithmetic processing device 56c in advance, and sequentially stores a value and an arithmetic result required in the arithmetic processing process. Next, the basic principle of the damping force control of each damping force variable shock absorber for the bounce motion suppression control executed in the present embodiment will be described.

First, when a damping force variable shock absorber having a damping force characteristic as shown in FIG. 8 is used, the gain characteristic of the output at which the vehicle body actually bounces with respect to the bounce motion input to act on the vehicle body is as follows. It appears as in FIG. Here, it should be noted that the horizontal axis sets the input vibration frequency to the vehicle body, that is, the frequency of the unsprung vertical velocity. Of these, relatively fast and large bounce movement, that is, unsprung vertical vibration in the middle and high frequency bands, reduces the input to the vehicle body by reducing the damping force to impair the occupant's riding comfort, while comparing Since the unsprung vibration in the low-frequency band, which is slow and slow, gives the occupant the heavy feeling seen in a vehicle with a large mass, it is sufficient to attenuate it so that the vibration does not fluffy and fluffy. Conceivable. In order to achieve this on the vehicle body side, that is, on the sprung side, considering the vibration input from the road surface, that is, the unsprung vertical vibration, the unsprung vertical vibration in the middle and high frequency bands generally has a vertical speed that is the inclination. From the viewpoint of energy saving, it is desirable to reduce the damping coefficient of the shock absorber and obtain the damping effect by making the damping force appearing as the product of the shock absorber and the direction of the finely fluctuating vibration smaller. On the other hand, the unsprung vertical vibration in the low frequency band generally has a small vertical speed, which is the slope. Therefore, the damping coefficient of the shock absorber is increased to increase the damping force that appears as the product of the two, and the It is desired to attenuate the unsprung vertical vibration. Then, in the input / output system of the vehicle bouncing bounce motion interposed with the bouncing motion suppression control system by the damping force variable shock absorber, the resonance frequency is set to a low frequency band of the sprung vertical speed, and the resonance frequency By reducing the gain from the state shown by the two-dot chain line in FIG. 10 to the state shown by the solid line, the gain for the sprung vertical velocity in the middle and high frequency bands to be aggressively attenuated is further reduced in the negative direction. While increasing the comfort, the solid feeling of the sprung vertical speed in the low frequency band can be changed from a fluffy one to a firm one. In order to achieve this with the damping force of the variable damping force shock absorber, that is, the damping coefficient characteristic, the unsprung vibration input in the low frequency band as described above is sufficiently large (high).
The damping coefficient can be set, and more importantly,
It is necessary to be able to set a sufficiently small (low) damping coefficient for an unsprung vibration input in a high frequency band.

In the input / output system of the bounce motion set as described above or the suppression control system thereof, in order to simply achieve the Karnopp's law, the swing input is used as shown by a two-dot chain line in FIG. The sprung vertical speed X _2i ′ (i = FL to R
For R), the target position may be set linearly with, for example, a proportional coefficient K. However, even when the vehicle is traveling on a good flat road surface, that is, even when a small vibration input is generated in a traveling state in which it is considered that there is no need to change and control the damping force, even if the vibration is within the soft range, Even if the damping force does not substantially change within the (SS range), rotating the stepping motor, that is, changing the position of the valve body, would be a waste of energy and would substantially reduce the stepping motor. There is also a problem of noise generated with rotation. Therefore, from the 'positive dead zone threshold X _2I0 against' sprung mass vertical velocity X _2i is a vibration input to a negative dead zone threshold (-X _2I0 ') and dead zone, sprung mass vertical velocity X _2i within the dead zone range ′, The bounce control target position P _DBi (i = _{FL to RR} ) is set to “0”, and when the sprung vertical speed X _2i ′ is not in this range, the sprung vertical speed X _2i ′ is increased. Accordingly, it is _assumed that the bounce control target position P _DBi increases linearly with the proportional coefficient K.

[0032] Here, sprung mass vertical velocity of the Figure 11 - Assuming the target position correlation characteristic control map, of all of the damping coefficient variable region, i.e., the extension side maximum position P _{MA X} from the compression side maximum position (- as all until P _MAX) can be applied to the bounce motion suppression control, the bounce control target position P _DBi (i = FL~RR) is when the extension side maximum position P _MAX, which corresponds to the bounce control target position P _DBi When 'the on vertical velocity X _2i extension side maximum spring' on vertical velocity X _2i spring and _MAX, bounce control target position sprung mass vertical velocity X _2i 'is the extension side maximum sprung mass vertical velocity X _2i' in _MAX or more areas P _DBi is fixed to the extension-side maximum position P _MAX . The bounce control target position P _DBi is the maximum pressure side position (−P _MAX ).
When the sprung vertical speed X _2i ′ corresponding to the bounce control target position P _DBi is the _compression- side maximum sprung vertical speed (−X _2i ′ _MAX ), the sprung vertical speed X _2i ′ is the compression-side maximum spring speed. The bounce control target position P _DBi is fixed at the _compression- side maximum position (-P _MAX ) in a region equal to or lower than the upper / lower speed (-X _2i ' _MAX ). Further, the sprung vertical velocity X _2i ′ when the bounce control target position P _DBi becomes the positive low damping threshold P _T1 is set to a positive low damping sprung vertical velocity threshold X _2i01 ′, and the negative low damping threshold P _{C1 is set.} The sprung vertical speed X _2i ′ at the time of is set as a negative low damping sprung vertical speed threshold (−X _2i01 ′).

As described above, the bounce control target position P for the sprung vertical speed X _2i ′ from the positive dead zone threshold X _2i0 ′ to the extension-side maximum sprung vertical speed X _2i ′ _MAX.
The slope K of the characteristic curve of _DBi (i = _{FL to RR} ) is expressed by the following equation.
_DBi is represented by the two equations below using a bounce control extension phase target position proportional coefficient alpha _2.

[0034] _{K = P MAX / (X 2i} 'MAX -X 2i0') ......... (1) P D = α 2 · P MAX = ((X 2i '-X 2i0') / (X 2i 'MAX - _{X 2i0 ')) · P MAX} ......... (2) Further, the negative dead zone threshold (-X _2I0') from the compression side maximum sprung mass vertical velocity (-X _{2i 'MA X)} sprung mass vertical velocity X _2i up'
Bounce control target position P _DBi (i = FL ~
Slope K of the characteristic curve of RR) is represented by the following three equations, also bounce control target position P _DBi at that time is represented by the following equation 4 using a bounce control compression phase target position proportional coefficient alpha _1.

K = (− P _MAX ) / (− X _2i ′ _MAX − (− X _2i0 ′)) (3) P _D = α ₁ · (−P _MAX ) = ((X _2i ′ − ( -X _2i0 ')) / (-X _2i ' _MAX -(-X _2i0 '))) (-P _MAX ) (4) Next, the roll motion suppression control in the suspension control device of the present embodiment is described. The basic principle of setting the damping coefficient of the variable damping force shock absorber for the above will be described while also considering the setting of the damping coefficient of the variable damping force shock absorber in the conventional roll motion suppression control.

First, in order to suppress the above-described conventional roll motion, the damping force generated by each damping force variable shock absorber is designed to increase as the roll speed increases. The roll control damping coefficient C of the absorber is set by the following five equations. C = C ₀ · R _R (5) Here, C ₀ in Equation 5 is a steady gain, and R _R is a roll angular velocity (roll rate).

Here, as shown in FIG.
Let us consider the case where _A temporally changes with time.
Aging of the roll angle R _A is, for example, for the case where the turning as lane change early reaches a maximum in a relatively short time, for example, the following aging of the roll angle R _A is relative to a time t It is assumed to be given by Equation 6. R _A = −A / ω · cosωt (6) where A in the expression 6 is the roll angular velocity amplitude, and ω is the roll angular frequency.

[0038] Thus, the roll angular velocity R _R is a differential value of the roll angle R _A is represented by the following Equation 7, the change over time appears as shown in Figure 12b. R _R = A · sinωt (7) By substituting the roll angular velocity R _R into the above equation (5), the roll control damping coefficient C is expressed by the following equation (8), and the change over time appears as shown in FIG. 12C.

C = C ₀ · A · sinωt (8) The damping force D / F expressed by each damping force variable shock absorber for the left and right wheels with respect to this damping coefficient C is the damping coefficient C
And the roll angular velocity R _R, and this damping force D / F is the tread width T _j between the left and right wheels (j = ForR, F is the front wheel side, R
Table as couple between showing) the rear wheel side, if indicated suppressing moment the roll movement generated in the vehicle body and the damping moment M _D of the roll control, the roll control damping moment M _D is the following formula (9) The change with time appears as shown by the solid line in FIG.

M _D = C · T _j · R _R = C · T _j · A · sin ωt = １ ／ · C ₀ · T _j · A ² · (1-cos 2ωt) (9) the attenuation coefficient of the conventional conventional shock absorber and the damping coefficient C ₁ of the constant value to the compression side to the extension side, the conventional shock absorber roll suppression by attenuating moment M _D ^* is expressed by the following equation (10), the change with time It appears as shown by the dashed line in FIG. 12d.

[0041] ^{_{M D * = C 1 · T}} j · A · sinωt ......... (10) conventional roll restraining represented by formula (9) damping moment M _D ^* with the conventional shock absorbers represented by the equation (10) controlled by as is apparent from the comparison between roll control damping moment M _D of the damping force variable shock absorber, in both phase and amplitude are changing. When the damping moment M _D of the damping force variable shock absorber according to the prior art roll control mathematically evaluated, for example, the attenuation coefficient of the conventional shock absorber C
If ₁ is equivalent to the constant gain C ₀ of the damping force variable shock absorber of the 5 formula, the damping moment M _D relative to the roll angular velocity R _R Out phase [pi, the phase [pi /
A phase delay occurs up to 2, a phase advance occurs up to the phase π, and the period becomes １ ／. Further, when the roll angular velocity amplitude A is "1" greater than, the greater the maximum amplitude of the damping moment M _D. That is, in a region where the damping moment M _D of the variable damping force shock absorber is smaller than the damping moment M _D ^* of the conventional shock absorber which is indicated by a broken line in FIG. 12d, the reaction force for suppressing the roll motion is insufficient. because so that the reaction force to suppress the roll motion becomes excessive in a region damping moment M _D is large attenuation moment M _D ^* variable damping force shock absorbers than the shock absorber, with a typical conventional shock absorber For a driver who is accustomed to the vehicle, an uncomfortable feeling such as a so-called roll stepping feeling is given.

In order to avoid such a problem, the extension side and compression side damping coefficients of the variable damping force shock absorber may be set to a certain high value at the same time as that of the conventional shock absorber. However, this cannot be achieved with the variable damping force shock absorber of FIG. Therefore, in the roll suppression control according to the present embodiment, in order to suppress the roll of the vehicle body from the direction of the roll motion acting on the vehicle, the damping force variable shock absorber of the wheel which should increase the compression coefficient on the compression side and the expansion coefficient on the extension side And the damping force of the variable shock absorber of the wheel for which the damping force of the wheel should be increased.
Is set to the relatively large predetermined compression side damping coefficient C _C2 or C _C3, and the damping force of the variable damping force shock absorber of the wheel to be increased in the expansion side is set to the relatively large predetermined expansion side damping coefficient C _T2 or Set to C _T3 .

Specifically, the front wheel side roll angular velocity R _RF and the rear wheel side roll angular velocity R _RR are calculated from the sprung vertical velocity X _2i ^′ according to the following equations (11) and (12). ^{_{R RF = (X 2FL '-X}} 2FR') / T F ......... (11) R RR = (X 2RL '-X 2RR') / T R ......... (12) thus defined is calculated Roll angular velocity R
By determining whether _Rj (j = ForR) is positive or negative, it is possible to determine at least the direction of the roll in a normal turn without a bounce input. That is, if each roll angular velocity R _Rj is positive, it means that a clockwise roll motion has occurred.
In the case of normal turning, it can be said that this is a left turning, and the pressure side damping coefficient of the damping force variable shock absorber of the turning outer wheel, that is, the right wheel, should be set to a large value.
That is, the extension-side damping coefficient of the variable damping force shock absorber for the left wheel may be set to be large. Conversely, if each roll angular velocity R _Rj is negative, it means that a counterclockwise roll motion has occurred, so it can be said that this is a right turn in a normal turn, and the turning outer wheel, that is, the left wheel, It is only necessary to set the compression side damping coefficient of the variable damping force shock absorber large, and it is sufficient to set the extension side damping coefficient of the variable damping force shock absorber of the inner turning wheel, that is, the right wheel. In this case, the damping force variable shock is adjusted so that the damping coefficient of the front wheel damping force variable shock absorber for which the front wheel is to be the turning outer wheel and the pressure side damping coefficient should be set to be large is the first predetermined pressure side damping coefficient C _C2. the valve body of the absorber, i.e. set the target position P _DRi for roll control of the step motor (i = FLorFR) to said first predetermined pressure side position value P _C2, a damping coefficient of extension side become similarly turning wheel The valve body of the variable damping force shock absorber, that is, the roll control target position P of the step motor, is set such that the damping coefficient of the variable damping force shock absorber of the front wheel to be set to be large becomes the first predetermined extension side damping coefficient C _T2.
Setting the _DRi to the first predetermined extension side position value P _T2.
In addition, the rear wheel is a turning outer wheel, and the variable damping force of the rear wheel should be set to be large so that the variable damping force of the rear wheel can be set to the second predetermined pressure side damping coefficient C _C3. Roll control target position P _DRi (i
= The RLorRR) is set to the second predetermined pressure side position value P _C3, likewise damping coefficient of the damping force control shock absorber rear wheels becomes the turning inner wheel to be set to a large damping coefficient of extension side is the second predetermined The valve body of the variable damping force shock absorber, that is, the roll control target position P _DRi of the stepping motor is set to the second predetermined extension side position value P _T3 so that the extension side damping coefficient C _T3 is obtained.

Each of the roll angular velocities R _Rj (j = For
Setting the roll control target position _PDRi until R) is smaller than a certain degree results in attenuating the sprung vertical speed in the low frequency band, and the smooth feeling of the solid feeling is obtained. Riding comfort is impaired, and it can be said that driving a step motor as an actuator to the roll control target position _PDRi wastes energy. The damping force can be varied until the sprung vertical speed X _2i ′ from which the roll angular velocity R _Rj is calculated is a road surface unevenness or a road undulation having different phases on the left and right wheels and the roll angular velocity R _Rj is small. If the damping coefficient of the shock absorber is controlled to follow, when the road surface input is lost, the damping force of the damping force variable shock absorber acts in the opposite direction, and the vehicle body may be jerky. Therefore, here is set the roll velocity dead zone threshold R _Rj0 the roll angular velocity R _Rj, the rolling angular velocity R _Rj only when it is the roll velocity dead zone threshold R _Rj0 above, roll control damping coefficient of the damping force control shock absorber Is set.

From the sum of the bounce control target position P _Bi (i = FL to RR) set as described above and the roll control target position P _Ri , the final value of the valve body of each damping force variable shock absorber is determined. The target position P _Di of the step motor is calculated. Since the control input used for both calculations is the sprung vertical speed X _2i ′, the target position P _Di output as the control amount is calculated.
As overlap in the bounce control component and the roll control component does not occur, (easy to understand considering the weighting coefficient) appropriate control gain in both K _B, the final by adding them from multiplying K _R The target position is P _Di.

However, in the bounce motion suppression control without considering the roll motion suppression control as described above, basically, as shown in FIG. 11, the entire damping force control range of the variable damping force shock absorber is set to the spring force which is the bounce motion input. Because it is set in the control variable range for the upper and lower speed X _2i ′,
Even if the bounce control target position P _Bi and the roll control target position P _Ri are weighted, there is a possibility that the sum of the two exceeds the positive / negative extension-side and pressure-side maximum positions P _MAX , (−P _MAX ). . In this case, when bounce vibration and roll vibration occur simultaneously, for example, when there is an upward bounce input to the left and right wheels of the vehicle in the same direction during a turn, for example, the variable damping force shock absorber on the turning outer wheel side Position P is the maximum pressure side position (-P
Continues to saturate the _MAX), but when the position P of the turning inner wheel side of the variable damping force shock absorbers are variably controlled until the extension side maximum position P _MAX, for example, a roll axis passing through the vehicle center of gravity where the aforementioned The vehicle body is vibrated in the roll direction so as to move in the vertical direction, and the vehicle behavior becomes unstable.

Therefore, in the present embodiment, the above problems are avoided by correcting the variable range of the bounce motion control amount with respect to the sprung vertical speed X _2i ′ when performing the roll motion suppression control. Specifically, in this embodiment, the roll control target position P _Ri of when the roll movement suppression control is executed, the first or second predetermined extension side position value P _T2, P _T3, or the first or second predetermined pressure side Since the position values P _C2 and P _C3 are set to constant values, the extension-side maximum position P _MAX excluding these values is set.
Alternatively, the range up to the compression-side maximum position (-P _MAX ) is defined as a bounce motion control amount variable region for the sprung vertical speed X _2i ′ as the bounce input. More specifically,
Substantial attenuation coefficient varies slightly from the attenuation coefficient as a target, the damping force and the position P, that the attenuation coefficient is the "0" to the extension side maximum position P _MAX or pressure side maximum position (-P _MAX) It is assumed that there is a linear relationship in the range, and in any case, any maximum position P _MAX
Or damping force (-P _MAX), i.e. the attenuation coefficients from the results in saturation, bounce motion control amount, excluding the extension side maximum position P _MAX or the compression side maximum position and (-P _MAX) of the roll control target position P _Ri the ratio between the variable region and bounce control correction coefficient _{C BR i (i = FL~RR)} , when calculating the final target position P _Di, the bounce control target position P this bounce control correction coefficient in the section _Bi C
_The bounce motion control amount variable region is corrected by multiplying _BRi .

That is, of the front left and right wheels, when a roll motion occurs due to turning or the like, the front turning outer wheel on the pressure side, that is, the bounce control correction coefficient C _{BRFC on the} front wheel pressure side is expressed by the following equation (13). The bounce control correction coefficient C _{BRFT for the} inner wheel, that is, the front wheel extension side, is given by the following equation (14). C _BRFC = (− P _MAX −P _C2 ) / (− P _MAX ) (13) C _BRFT = (P _MAX −P _T2 ) / P _MAX (14) Here, FIG. As described above, the absolute value | P _C2 | of the first predetermined compression-side position value is larger than the absolute value | P _T2 | of the first predetermined extension-side position value, and the absolute value | P _MAX | Absolute value of position |
Is larger than −P _MAX |, the calculated front wheel pressure side bounce control correction coefficient C _BRFC is larger than the front wheel extension side bounce control correction coefficient C _BRFT . Further, the roll control target position P _DRi by the roll motion suppression control is:
Absolute value | R _RF | of front wheel side roll speed is dead zone threshold value R _RF0
Otherwise, the front-wheel-side bounce control correction coefficient C _BRF for the absolute value of the front-wheel-side roll speed | R _RF | appears as shown in FIG. 13A. That is, in the region where the absolute value | R _RF | of the front wheel side roll speed is smaller than the dead zone threshold value R _RF0 , both the front wheel side bounce control correction coefficient _CBRF is set to “1”, and the absolute value | R of the front wheel side roll speed is set.
_{When RF} | is equal to or greater than the dead zone threshold value _RRF0 , the front-side bounce control correction coefficient C _{BRF on the compression} side is set to the correction coefficient C _BRFC by the above equation (13), and the front-wheel bounce control correction coefficient C _BRF on the extension side is the above equation (14). _Is set to the correction coefficient C _BRFT .

On the other hand, of the rear left and right wheels, the bounce control correction coefficient C _{BRRC for the} rear turning outer wheel on the pressure side, that is, the rear wheel pressure side, when a roll motion occurs due to turning, etc. The bounce control correction coefficient C _{BRRT for the} turning inner wheel, that is, the rear wheel extension side, is given by the following equation (16). C _BRRC = (− P _MAX −P _C3 ) / (− P _MAX ) (15) C _BRRT = (P _MAX −P _T3 ) / P _MAX (16) Here, FIG. As described above, the absolute value | P _C3 | of the second predetermined compression side position value is larger than the absolute value | P _T3 | of the second predetermined extension side position value, and the absolute value | P _MAX | of the maximum extension side position is the compression side maximum. Absolute value of position |
−P _MAX |, the calculated rear wheel pressure side bounce control correction coefficient C _BRRC is larger than the rear wheel extension side bounce control correction coefficient C _BRRT . Further, the roll control target position P _DRi by the roll motion suppression control is:
The absolute value | R _RR | of the rear wheel side roll speed is the dead zone threshold value R _RR0
Otherwise, the rear wheel bounce control correction coefficient C _BRR for the absolute value | R _RR | of the rear wheel roll speed appears as shown in FIG. 13B. That is, in a region where the absolute value | R _RR | of the rear wheel side roll speed is smaller than the dead zone threshold value R _RR0 , both the rear wheel side bounce control correction coefficient C _BRR is set to “1”, and the absolute value of the rear wheel side roll speed is set. Value | R
_{When RR} | is _{equal to} or greater than the dead zone threshold value R _{RR0, the compression-} side rear wheel-side bounce control correction coefficient C _BRR is set to the correction coefficient C _{BRRC according} to the equation (15), and the extension-side rear wheel-side bounce control correction coefficient C _BRR is set to It is set to the correction coefficient C _BRRT by equation (16).

Each of these correction coefficients C _BRFC , C _BRFT , C
_BRRC and C _BRRT are suffixes i (= FL) representing the corresponding wheels, respectively.
Since denoted bounce control correction coefficient C _BRi with ~RR), the final target position P _Di is given by the following 17 formula. _{P Di = C BRi · K B} · P DBi + K R · P DRi ......... (17) So then, at the same time capable body attitude change suppressing control the bounce motion suppression control and the roll motion suppression control in accordance with the basic principle For this purpose, FIG. 15 shows the arithmetic processing of damping force control executed by the arithmetic processing unit 56c of the microcomputer 56. In the present embodiment is calculated according to calculation equation rather than a map search for bounce control target position P _{DB i} is set to the basic. This calculation process is executed by a timer interrupt at every predetermined sampling time as described later, and the final target position _PDi calculated here and the damping force D / F at the final target position _PDi are achieved. For example, the step amount S _i or the like is calculated by, for example, the damping force variable shock absorber 3F of the front left wheel.
L, damping force variable shock absorber 3FR for front right wheel, variable damping force shock absorber 3RL for rear left wheel, variable shock absorber 3RR for rear right wheel I do. The data and information necessary for the current such positions P _A stored updated in the storage device 56d for each calculation process, even without exceptional output processing step, the arithmetic processing unit 56c for each said operation processing It shall be read into a buffer or the like at any time.

That is, the process of FIG. 15 is executed as a timer interrupt process at predetermined time intervals ΔT (for example, 3.3 msec). First, at step S1, each of the vertical acceleration sensors 51FL to 51FL is detected.
Each spring vertical acceleration detection value X _2i ″ detected by RR (i = FL
~ RR). Next, the process proceeds to step S2, in which each of the sprung vertical acceleration detection values X _2i ″ read in step S1 is subjected to a high-pass filter process using a digital high-pass filter constructed by a program, for example. The drift superimposed component of the upper / lower acceleration detection value X _2i ″ is removed. The cutoff frequency of the digital high-pass filter can be set by appropriately selecting a temporary variable of a program for constructing the filter, as is known.

Next, the process proceeds to step S3, in which a low-pass filter is applied to each of the sprung vertical acceleration detection values X _2i ″ from which the drift superimposed component has been removed in step S2 by, for example, a digital low-pass filter constructed by a program. Then, the sprung vertical velocity X _2i ′ phase-matched is calculated as an integral value of the digital signal.
As is known, a temporary variable of a program for constructing the filter can be appropriately selected and set. Further, the calculation of each sprung vertical speed X _2i ′ can be performed not by the low-pass filter processing but by the existing integral calculation processing.

Next, the process proceeds to step S4, where it is determined whether or not the sprung vertical speed X _2i ′ is smaller than “0”, that is, whether or not the sprung vertical speed X _2i ′ is negative. To step S5, otherwise, to step S5.
Move to 6. At the step S5, the compression side maximum sprung mass vertical velocity (-X _{2i 'MAX)} and the on each spring calculated set in step S3 vertical velocity X _2i' and preset the negative sprung mass vertical velocity dead zone threshold Using (−X _2i0 ′), the bounce control pressure side target position proportional coefficient α ₁ is calculated based on the above equation (4), and then the _{process proceeds} to step S7. Note that the absolute value of the negative dead zone threshold value | (−
X _2i0 ') | is the absolute value of the positive dead zone threshold | X _2i0 ' |
When it is equivalent to the above, double parentheses in the four equations may be removed.

In step S7, step S5
In bounce control compression phase target position proportional coefficient alpha ₁ is calculated to determine whether a "1" or more, if the bounce control compression phase target position proportional coefficient alpha ₁ is "1" or more in step S8 The process proceeds to step S9 otherwise. In the step S8, the bounce control pressure side target position proportional coefficient α _{1 is set} to “1”.
Then, the process proceeds to step S9.

In step S9, step S5
In the calculated bounce control compression phase target position proportional coefficient alpha ₁ is "0" determines whether less or is, if the bounce control compression phase target position proportional coefficient alpha ₁ is "0" or less in step S10 The process proceeds to step S11 otherwise. In the step S10, the transition from set to "0" to the compression side target position proportional coefficient alpha ₁ in the step S11.

At the step S11, at the step S11
5 or step S8, or the pressure side in accordance with the four equation using the set bounce control compression phase target position proportional coefficient alpha ₁ and the pressure side maximum position (-P _MAX) at step S10, i.e., the negative direction of the bounce control target position P
_After calculating _DBi , the _{process proceeds} to step S12. on the other hand,
In step S6, the extension-side maximum sprung vertical speed X
_2i ′ _MAX and the respective sprung vertical speed X _2i ′ calculated and set in the step S3 and the predetermined positive sprung vertical speed dead zone threshold X _2i0 ′, using the bounce control extension side based on the above equation (2). After calculating the target position proportional coefficient α ₂ , the process proceeds to step S13.

In the step S13, the step S
Bounce control extension phase target position proportional coefficient alpha ₂ calculated in 6 determines whether a "1" or more, if the bounce control extension phase target position proportional coefficient alpha ₂ is "1" or more The process proceeds to step S14, and if not, the process proceeds to step S15. In step S14, the bounce control extension side target position proportional coefficient α ₂
Is set to "1", and then the process proceeds to step S15.

At the step S15, at the step S15
Bounce control extension phase target position proportional coefficient alpha ₂ calculated in 6 is equal to or less than "0", if the bounce control extension phase target position proportional coefficient alpha ₂ is equal to or less than "0" The process proceeds to step S16, and if not, the process proceeds to step S17. In step S16, the bounce control extension side target position proportional coefficient α ₂
Is set to "0", and the routine goes to the step S17.

At the step S17, at the step S17
Using the bounce control expansion side target position proportional coefficient α ₂ and the expansion side maximum position P _MAX set in step 6 or step S14 or step S16, the expansion side, that is, the forward direction bounce control target position P _DBi is _calculated according to the above equation (2 ₎ . After the calculation, the process proceeds to step S12. Then, in step S12, the vertical sprung calculated in the step S3 velocity X _2i 'and front wheel tread width T _F or the rear wheel side tread width T using said _R 11 Formula or 12 front and rear wheel side roll angular velocity according to equation After calculating and setting R _Rj (j = ForR), the process proceeds to step S18.

At the step S18, at the step S18
Absolute value of the front wheel side roll angular velocity calculated in step 12 | R _RF |
Is greater than or equal to the front wheel side roll angular velocity dead band threshold _RRF0 , and the absolute value | R of the front wheel side roll angular velocity is determined.
_{If RF} | is equal to or greater than the front wheel roll angular velocity dead zone threshold value R _RF0 , the process proceeds to step S19; otherwise, the process proceeds to step S20.

In step S20, the roll control target positions P _DRi (i = FLorFR) of the front left and right wheels are both set to “0”, and then the flow shifts to step S21. In step S21, the bounce control correction coefficient C _BRi (i = FLorFR) for the front left and right wheels is both set to “1”, and the process proceeds to step S22. On the other hand, in step S19, it is determined whether or not the front wheel side roll angular velocity R _RF calculated in step S12 is greater than “0”, that is, whether or not the front wheel side roll angular velocity R _RF is positive. If so, the process proceeds to step S23; otherwise, the process proceeds to step S24.

In step S23, the front left
Roll control target position P for right wheel _DRi(I = FLorFR)
Of the front left wheel roll system
Your target position P_DRFLIs the first predetermined extension side position
Value P_T2And the roll motion
Front right wheel roll control target position P_DRFRThe said
1 Predetermined pressure side position value P_C2Step S after setting to
It moves to 25.

In step S25, of the front left and right wheel bounce control correction coefficients C _BRi (i = FLorFR), the front left wheel bounce control correction coefficient C
_BRFL is set to the front wheel bounce-side bounce control correction coefficient C _BRFT given by the above equation (14), and the front right wheel bounce control correction coefficient C _BRFR which becomes the _compression side by the roll motion is given by the front wheel pressure-side bounce control correction given by the above equation (13). Coefficient C
_After setting to _BRFC , the process _proceeds to step S22.

On the other hand, in step S24, among the roll control target positions P _DRi (i = FLorFR) of the front left and right wheels, the front left wheel roll control target position P _DRFL which becomes the pressure side by the roll motion is set to the first predetermined pressure side position value. P
And sets the _C2, step S26 the front right wheel roll control target position P _DRFR as the extension side by the rolling motion after setting the first predetermined extension side position value P _T2
Move to

In step S26, of the front left and right wheel bounce control correction coefficients C _BRi (i = FLorFR), the front left wheel bounce control correction coefficient C
_BRFL is set to the front wheel pressure side bounce control correction coefficient C _BRFC given by the above equation (13), and the front right wheel bounce control correction coefficient C _BRFR to be _extended by the roll motion is given by the front wheel bounce control given by the above equation (14). Correction coefficient C
_After setting to _BRFT , the _{process proceeds} to step S22.

At the step S22, at the step S22
Absolute value of rear wheel side roll angular velocity calculated in step 12 | R _RR |
_Is greater than or _{equal to the} rear wheel side roll angular velocity dead zone threshold value R _RR0 , and the absolute value | R of the rear wheel side roll angular velocity is determined.
_{If RR} | is _{equal to} or greater than the rear wheel roll angular velocity dead zone threshold value R _RR0 , the _{process proceeds} to step S27; otherwise, the _{process proceeds} to step S28.

In step S28, the roll control target positions P _DRi (i = RLorRR) of the rear left and right wheels are both set to “0”, and then the flow shifts to step S29. In step S29, the bounce control correction coefficients C _BRi (i = RLorRR) for the rear left and right wheels are both set to “1”, and the process proceeds to step S30. On the other hand, in step S27, it is determined whether or not the rear wheel side roll angular velocity R _RR calculated in step S12 is greater than “0”, that is, whether or not the rear wheel side roll angular velocity R _RR is positive. Is positive, the process proceeds to step S31; otherwise, the process proceeds to step S32.

In step S31, the rear left
Roll control target position P for right wheel _DRi(I = RLorRR)
Of the rear left wheel roll system
Your target position P_DRRLIs the second predetermined extension side position
Value P_T3And the roll motion
Rear right wheel roll control target position P_DRRRThe said
2 Predetermined pressure side position value P_C3Step S after setting to
Move to 33.

In step S33, the rear left wheel bounce control correction coefficient C _BRi (i = RLorRR) of the rear left wheel bounce control, which becomes the extension side due to the roll motion, is determined.
_BRRL is set to the rear wheel bounce-side bounce control correction coefficient C _BRRT given by the above equation (16), and the rear right wheel bounce control correction coefficient C _BRRR, which becomes the _compression side by the roll motion, is given by the rear wheel bounce given by the equation (15). Control correction coefficient C
_After setting to _BRRC , the _{process proceeds} to step S30.

On the other hand, in step S32, of the rear left and right wheel roll control target positions P _DRi (i = RLorRR), the rear left wheel roll control target position P _DRRL which becomes the pressure side due to the roll motion is _changed to the second predetermined pressure side position value. P
And sets to _C3, step S34 the right wheel roll control target position P _DRRR after the extension side by the rolling motion after setting the second predetermined extension side position value P _T3
Move to

In step S34, of the bounce control correction coefficients C _BRi (i = RLorRR) for the rear left and right wheels, the rear left wheel bounce control correction coefficient C
_BRRL is set to the rear wheel pressure side bounce control correction coefficient C _BRRC given by the above equation (15), and the rear right wheel bounce control correction coefficient C _BRRR to be _extended by the roll motion is set to the rear wheel extension side given by the above equation (16). Bounce control correction coefficient C
_After setting to _BRRT , the _{process proceeds} to step S30.

In step S30, the bounce control target position P calculated and set in each of the above steps is set.
_DBi (i = _FL~RR), roll control target position P _DRi, using bounce control correction coefficient C _BRi and bouncing control gain K _B and roll control gain K _R, a final target position P _Di in accordance with the 17 formula After the calculation and setting, the process proceeds to step S35.

At the step S35, at the step S35
Target position P calculated in 30 _DiFrom the above
Current position P updated and stored in the storage device 56d_Ai
To reduce the rotation angle of the step motor by the step amount S._iage
Then, the process proceeds to step S36. The step
In step S36, the values calculated and set in step S35 are set.
Absolute value of step amount | S_jIs |
Maximum step amount S achieved by arithmetic processing_MAXIs less than
Is determined, the absolute value of the step amount | S_j|
Large step amount S_MAXIf not, step S37
Otherwise, to step S38.
You.

At the step S37, at the step S37
Step amount S calculated and set at 35 _jStep
Step amount S which is a control signal to the motor_jSet to
Then, control goes to a step S39. In the step S38
Is the step amount S calculated and set in step S35.
_jIs greater than "0", that is, whether it is positive.
And the step amount S _jStep if is positive
Move to S40, otherwise go to step S41
Transition.

In step S40, the step amount _Sj , which is a control signal to the step motor, is set to the positive value _SMAX of the maximum step amount, and then the flow shifts to step S39. In the step S41, the transition from to set the step amount S _j is a control signal to the step motor to a negative value (-S _MAX) of the maximum stepping amount to the step S39.

At the step S39, at the step S39
Step amount S set in any of 37, S40 and S41
_jAs a control signal to the step motor,
Output to the dynamic circuit 59FL-59RR
Returns to gram. Next, the vehicle according to the calculation processing of FIG.
The operation of the body posture change suppression control will be described. here
First, the sprung vertical speed X between the left and right wheels_2i'Has small deviation
As a result, in step S12 of the arithmetic processing of FIG.
Absolute value of each roll angular velocity | R _Rj|
Dead zone threshold R_Rj0Because of the smaller
Front left and right wheel roll control target position in step S20
P_DRFL, P_DRFRAre both set to “0” and the
In step S21, the front left and right wheel bounce control correction coefficient C_BRFL,
C_BRFRAre also set to "1", and at step S28
Rear left and right wheel roll control target position P_DRRL, P_DRRRIs
Is set to "0", and in step S29,
Right wheel bounce control correction coefficient C_BRRL, C_BRRRBoth are “1”
Is set, and as a result, the eye calculated in step S30 is set.
Mark position P_DiIs the bounce control target position P_DBi
Bounced by the calculation process
The operation of the exercise suppression control will be described.

As shown in FIG. 11, assuming that the position P is set linearly with respect to the sprung vertical velocity X _2i ′ excluding the dead zone in steps S5 to S17 in this arithmetic processing, as shown in FIG. The damping force characteristic shown appears for the sprung vertical speed X _2i ′ as shown in FIG. 15C. That is, the scale of the position-damping force characteristic shown in FIG.
Assuming that the scale of the sprung vertical speed-damping force characteristic shown in FIG. 15C is equivalent to the soft range (SS range) of the sprung vertical speed-damping force characteristic shown in FIG.
It is sufficient to consider that the expansion-side control range (HS range) and the compression-side control range (SH range) are located outside the sprung vertical speed dead zone which is maintained at.
For this sprung vertical velocity-damping force characteristic, the sprung vertical velocity X _2i ′ as shown in FIG. 15A is used as a transient vibration input, that is, as a slow and relatively large vibration input in the relatively low frequency band described above. Let's consider the effect of input.

First, as an initial input, a bar that increases in the positive region
Vertical speed X_2i'Is time t_TenWith the positive low damping sprung
Vertical speed threshold X_2i01'' And continue to increase,
Eventually, the increase in the extension side damping force described later
The increase slope becomes smaller, and at a certain time,
Begins to decrease in the region, and eventually at time t₂₀At the positive low attenuation
Upper and lower speed threshold X_2i01' On the contrary,
Absolute value of sprung vertical velocity when passing through the SS range
｜ X_2i'| Is small and the SS range including the dead band range
Because the damping coefficient on the extension side and compression side set in the box is small
In addition, the variable damping force shock absorber in the SS range
Damping force D / F achieved in the minimum damping force D / Fmin
It is considered that the
The damping force D / F achieved at the position P
Sprung vertical speed X excluding the dead zone range_2i'And the linear function
At the time t_TenFrom time t₂₀Until
Interval t _Ten~ T₂₀Now, steps S6 to S6 of the arithmetic processing in FIG.
In step S17, the sprung vertical speed X_2i'Synchronized with the increase or decrease
Bounce control target on the extension side to achieve improved damping coefficient
Position P_DBiIs calculated and set to the target
Condition P_DiIs the bounce control target position P_DBionly
, This bounce control target position
P_DBiIn particular, the extension side damping force D / F corresponding to
Occurs as shown. Conversely speaking, sprung vertical speed
X_2i'Is the effect of the extension damping force D / F according to the increase / decrease
Is attenuated. This means that the body and wheels
Work as a force to hinder the movement
Direction movement can be reduced. At this time, the pressure side decreases
The damping force D / F is minimum as shown in FIG.
From the top of the road, the road is moved upwards
Has almost no effect on the vehicle body.

The sprung vertical velocity X _2i ′, which continues to decrease, begins to decrease in the negative region, and at time t ₃₀ falls below the negative low damping sprung vertical velocity threshold (−X _2i01 ′). Continue but eventually gradually decreasing slope decreases its by the compression side damping force increasing effect to be described later, beyond a minimum point at a certain time begins to increase at the negative region, eventually the negative vertical on low attenuation spring at time t _{40 Exceeded the} speed threshold ( _-X2i01 '). Time t ₃₀ from the time t ₃₀ to time t ₄₀ ~t ₄₀
Now, steps S5 to S1 of the arithmetic processing in FIG.
The bounce control target position P mainly on the compression side which achieves a damping coefficient synchronized with the increase and decrease of the sprung vertical speed X _2i ′ at 1.
_DBi is calculated and set, corresponding to the bounce control target position P _DBi, especially since the compression side damping force D / F is generated As shown in Figure 15b, the sprung mass vertical velocity X _2i '
Is effectively attenuated by the damping force D / F according to the increase / decrease of the self. That is, contrary to the case where the vehicle body and the wheel are separated from each other, the damping force is generated only when the two approach each other, and as a result, the downward movement of the vehicle body can be reduced. Further, at this time, since the extension side damping force D / F is minimized as shown in FIG. 12, the influence on the vehicle body does not occur even if the wheels are moved downward due to the unevenness of the road surface. Almost no. Note that the absolute value | X _2i '| of the sprung vertical velocity at the minimum point is smaller than the absolute value | X _2i ' | of the sprung vertical velocity at the maximum point.

Since such a relatively slow and large vibration input tries to vibrate the vehicle body alternately in the vertical direction, the same applies to the time t _{40 (} not shown) after the time t _40. The value of the sprung vertical speed X _2i ′ is repeatedly increased and decreased while the value gradually converges, and at least the maximum value and the minimum value of the sprung vertical speed X _2i ′ are both the negative low damping sprung vertical speed threshold value (−
_X2i01 ') to the positive low damping sprung vertical speed upper limit threshold X
'Until converges to S-S range up, on the spring vertical velocity _{X 2i'} 2i01 damping force D / F of the damping force variable shock absorber in response to an increase or a decrease of is variably controlled.

As described above, since the effective damping force D / F is exhibited in each damping force variable shock absorber with respect to the large sprung vertical speed X _2i ′ even at a relatively low frequency, the vehicle body swings due to this. Such a change in posture is quickly attenuated and converged, giving the occupant a firm feeling. on the other hand,
At a low frequency and a small sprung vertical speed X _2i ′, the damping force D / F of each damping force variable shock absorber is equal to the _SS
Although the vehicle converges to the minimum damping force D / Fmin in the range, the vehicle body swings to such an extent that it does not flutter, but it provides a smooth ride comfort such as the solid feeling in the low-speed running state as described above. . Further, even at a high frequency and a small sprung vertical speed X _2i ′, the damping force D / F of each damping force variable shock absorber converges to the minimum damping force D / Fmin in the SS range. In such a case where the fine road surface irregularities are continuous, the small sprung vertical velocity X _2i ′ at such a high frequency generally has a large change rate, and thus changes finely and quickly with a small damping coefficient corresponding to the minimum damping force D / Fmin. The damping force that appears as the product of the sprung vertical speed X _2i ′ effectively attenuates the road surface input in accordance with the direction and magnitude of the road surface input, and provides a good ride comfort.

On the other hand, when the vehicle is turned on the right road, which is extremely flat and has almost no road surface unevenness or road undulation, for example, the lateral acceleration and the cornering force are used as couples, and a counterclockwise rotation is made with respect to the roll axis passing through the center of gravity of the vehicle. Roll motion (roll moment) occurs. Due to the roll moment, an upward acceleration is generated on the vehicle body, that is, on the right side of the spring, and at the same time, a downward acceleration is generated on the left side of the spring. Therefore, for each sampling time ΔT during which the calculation processing of FIG. 14 is executed, of the sprung vertical acceleration detection values X _2i ″ read in step S1 of the calculation processing, the front-rear right sprung vertical acceleration detection value X _2FR ", X _2RR" becomes positive upward, the front and rear left spring vertical acceleration detection value X _{2F L} ", X _2RL"
Is downward and negative. Therefore, among the sprung vertical speeds X _2i ′ calculated in step S3 of the same calculation process, the front and rear right sprung vertical speeds X _2FR ′ and X _2RR ′ become positive _values and the front and rear left sprung vertical speeds X _2FL ′ and X _2FL ′. X _2RL 'is considered to be a negative value. The sprung vertical speed X _2i ′ calculated in this manner.
On the other hand, in steps S5 to S17 of the calculation processing in FIG. 14, the bounce control target position P _DBi is calculated and set in the same manner as described above.
The bouncing control gain K _B in said 17 Expressions 0,
It is considered that the effect does not appear on the damping coefficient achieved by the damping force variable shock absorber or the resulting damping force.

When the sprung vertical speed X _2i ′ is generated by the counterclockwise roll motion as described above, the front and rear wheel-side roll angular speed R _Rj (j = ForR) calculated in step S12 of the calculation processing in FIG. _Are both negative values and the absolute value is _{equal to} or greater than the front and rear wheel side roll angular velocity dead zone threshold value R _Rj0 , and from step S18 to step S
19 through the processing proceeds to step S24, where the front left wheel roll control target position P _DRFL is set to the first predetermined pressure side position P _C2, front right wheel roll control target position P _DRFR the first predetermined extension side position P Set to _T2 ,
In the following step S26, the front left wheel bounce control correction coefficient C
_BRFL is set to the front wheel pressure side correction coefficient C _BRFC , the front right wheel bounce control correction coefficient C _BRFR is set to the front wheel expansion side correction coefficient C _BRFT , and then the process proceeds from step S22 through step S27 to step S32, where the rear The left wheel roll control target position P _DRRL is _equal to the second predetermined pressure side position P _C3.
It is set to the rear right wheel roll control target position P _DRRR is set to the second predetermined extension side position P _T3, rear left wheel bounce control correction coefficient C _BRRL in following S23 is set to the rear wheel compression side correction coefficient C _BRRC , The rear right wheel bounce control correction coefficient C _BRRR is set to the rear wheel extension correction coefficient C _BRRT , and the final target position P _Di is calculated in step S30 using these set values and calculated values, and step S35 ~
The position P of the stepping motor of each variable damping force shock absorber changes to the target position _PDi according to the control signal step amount _Sj output in step S41.

[0084] Here, when calculating the target position P _Di to be executed by the operation given in step S30 of FIG. 14, by the bounce control gain K _B and bounce control correction coefficient C _BRi, wherein step S4~ step S1
When the influence of the bouncing control target position P _DBi calculated deems not appear in the target position P _Di in 7, 16
As shown in c, the damping coefficient C _FL of the front left wheel damping force variable shock absorber among the respective damping force variable shock absorbers becomes the first predetermined pressure side damping coefficient C _C2 , and the damping force of the front right wheel damping force variable shock absorber is reduced. The coefficient C _FR becomes the first predetermined extension side damping coefficient C _T2 , the damping coefficient C _RL of the rear left wheel damping force variable shock absorber becomes the second predetermined pressure side damping coefficient C _C3 , and the rear right wheel damping force variable shock absorber the damping coefficient C _RR is supposed to be the second predetermined extension side damping coefficient C _T3. Therefore, given the same rolling motion and the 12, the damping force D / F of each damping force variable shock absorbers of the left and right wheels obtained by multiplying the roll angular velocity R _Rj to the attenuation coefficient C _i is between the left and right wheels tread Width T
_{j (j} = ForR, F is the front wheel side, R represents shows the rear wheel side) Table will couple between, the resulting vehicle body occurs suppressing roll control damping the rolling motion moment M _D below 20 formula The change over time appears as shown by the solid line in FIG. 16d.

M _D = C _i · T _j · R _R = C _i · T _j · A · sinωt (20) where C _i = C _C2 or C _T2 or _{C C3} or _{C T3} Damping coefficient C ₁ with a constant damping coefficient
Since the is the roll suppressing damping moment M _D ^* is the conventional conventional shock absorber is as the equation (10), each damping coefficient C _i of the damping force variable shock absorber is achieved by the rolling motion suppression control, the conventional Shock absorber damping coefficient C ₁
By equal or substantially equal, it will be able coincident or almost coincident with the 20 equations of the damping moment M _D and the damping moment M _D ^*, to the driver accustomed to ride on a vehicle having a conventional shock absorber and There is no discomfort.

Conversely, if the vehicle turns left on the good road,
Roll motion (roll moment) occurs around
Then, a downward acceleration occurs on the vehicle body, that is, on the right side on the spring.
At the same time, upward acceleration occurs on the left side of the spring
The sampling from which the arithmetic processing of FIG.
It is read at step S1 of the arithmetic processing at every time ΔT.
Sprung vertical acceleration detection value X_2i″, Front and rear right sprung
Vertical acceleration detection value X_{2F R}″, X_2RR″ Is downward and negative
And the vertical acceleration detection value X_2FL″,
X_2RL″ Is a positive value in the upward direction, and is calculated in step S3.
The sprung vertical speed X _2i', The front and rear right sprung vertical speed
Degree X_2FR', X_2RR'Is a negative value.
Speed X_2FL', X_2RL'Is considered positive,
Front and rear wheel side roll angular velocity R calculated in step S12_Rj(J
= ForR) are both positive values, and their absolute values are on the front and rear wheel sides.
Roll angular velocity dead zone threshold R_Rj0Given that
The process proceeds from step S18 to step S19 of the arithmetic processing.
The process proceeds to step S23 where the front left wheel roll control target position is determined.
Condition P_DRFLIs the first predetermined extension side position P_T2Set in
And the front right wheel roll control target position P_DRFRIs
First predetermined pressure side position P _C2Set to and follow steps
In S25, the front left wheel bounce control correction coefficient C_BRFLIs the front wheel extension side
Correction coefficient C_BRFTThe front right wheel bounce control correction
Number C_BRFRIs the front wheel pressure side correction coefficient C_BRFCIs set to
From step S22 to step S3 through step S27
1 where the rear left wheel roll control target position P
_DRRLIs the second predetermined extension side position P_T3Is set to
Right wheel roll control target position P_DRRRIs the second predetermined pressure
Side position P_C3Is set in the subsequent step S33.
Left wheel bounce control correction coefficient C_BRRLIs the rear wheel extension correction coefficient C
_BRRTAnd the rear right wheel bounce control correction coefficient C_BRRRIs
Rear wheel pressure side correction coefficient C_BRRCAre set to each of these settings
And the calculated target value in step S30.
P_DiAre calculated, and steps S35 to S41 are calculated.
Control signal step amount S output at_jBy each damping force
Position of the step motor of the variable shock absorber
P is this target position P_DiChanges to

At this time, _{assuming that} the influence of the bounce control target position P _DBi does not appear in the final target position P _Di , the damping coefficient achieved by each damping force variable shock absorber or the generated damping force is equal to the right damping force. The left and right are interchanged with each other when turning, and as a result, the direction of action of the damping moment is only changed between left and right, and the other operation and effects are the same or almost the same as during the right turn. .

When the bouncing motion occurs while the roll motion is occurring, such as during the turning, the bounce control for actively suppressing the bouncing motion is performed in steps S5 to S17 in FIG. target position P _DBi is calculated and set, it controls the gain from added to the final target position P _Di via a K _B, the damping force control range of the variable damping force shock absorber using the entire region, the bounce motion Are also effectively attenuated and converged. Of course, since the damping force control range applied to the bounce motion suppression control is corrected by the bounce control correction coefficient C _BRi , for example, the position P of the variable damping force shock absorber on the turning outer wheel is set to the _compression- side maximum position (−P
_MAX ), but the position P of the variable damping force shock absorber on the turning inner wheel side is variably controlled up to the maximum extension position P _MAX, so that the vehicle body is vibrated in the roll direction. The vehicle behavior does not become unstable.

As described above, each of the vertical acceleration sensors 51FL
14 to 51RR and steps S1 to S3 of the calculation processing of FIG. 14 correspond to the vehicle body movement state detecting means of the suspension control device of the present invention, and similarly, step S12 of the calculation processing of FIG. FIG. 1
Steps S18 to S34 of the calculation processing of Step 4 correspond to the roll motion control means, and the entire calculation processing of the controller 4 and FIG. 14 corresponds to the control means.
RR corresponds to the actuator, and the damping force variable shock absorbers 3FL to RR correspond to the damping force variable shock absorbers.

In the above embodiment, the roll angular velocity as the roll motion state is detected or calculated from the deviation of the sprung vertical velocity in the vicinity of the left and right wheels. There is no need to provide it, and the roll angular acceleration or the roll angle itself may be used, and the detection means may be directly detected using a roll rate gyro, a roll angular acceleration sensor, or the like. However, when the damping force is evaluated in the form of a product with the damping coefficient as described above, it is needless to say that it is necessary to temporarily convert these to the roll angular velocity.

Further, in the above embodiment, only the case where the bounce motion suppression control and the roll motion suppression control are executed as the vehicle body posture change suppression control is described in detail. It is not an essential requirement, and only the roll motion suppression control may be performed as described above. Further, a control mode for detecting the pitch motion as a motion state caused by a change in the posture of the vehicle body and suppressing the movement may be incorporated in the vehicle body posture change suppression control as needed.

Further, in the above embodiment, the case where the valve element 31 for controlling the damping force is formed in a rotary type has been described. However, the present invention is not limited to this. Different flow paths may be formed on the side and in this case, in this case, the step motors 41FL to 4FL
A pinion may be connected to the rotation shaft 41a of the 1RR, and a rack that meshes with the pinion may be attached to the connection rod 42 or an electromagnetic solenoid may be used to control the sliding position of the valve element 31.

Further, in the above-described embodiment, the case where the skyhook approximation control in which the vertical acceleration of the vehicle body is detected and the damping force is controlled based on the detected acceleration is described, is not limited thereto. Instead, a stroke sensor that detects the relative displacement between the vehicle body and the wheels is installed separately,
Based on the relative speed _XDi 'obtained by differentiating the relative displacement detection value _XDi of the stroke sensor and the above-described vehicle body vertical speed _X2i ', a calculation of the following equation 10 is performed to calculate a damping coefficient C. For example, the target position may be calculated with reference to a map corresponding to FIG.

C = C _{S ×} X _2i ′ / X _Di ′ (10) where C _S is a preset damper damping coefficient.
In the above embodiment, the microcomputer 5
6 has been described, but the present invention is not limited to this, and may be configured by combining electronic circuits such as arithmetic circuits.

In the above embodiment, the vertical acceleration sensors 51FL to 51RR are located at the wheels 1FL to 1RR of the vehicle body 2.
Is described, but one of the vertical acceleration sensors may be omitted, and the vertical acceleration at the omitted position may be estimated from the values of the other vertical acceleration sensors. In the above embodiment, the step motor 41
The case where open loop control is performed on FL to 41RR has been described. However, the present invention is not limited to this, and closed loop control may be performed by detecting the rotation angle of a step motor with an encoder or the like and feeding it back.

[0096]

As described above, according to the suspension control apparatus of the present invention, the state of roll motion generated on the vehicle body is detected, and the damping force variable on the compression side is determined according to the roll motion state, that is, the direction of the roll moment. configuration and set to a relatively large constant predetermined extension side damping coefficient extension side damping coefficient of the damping force control shock absorber comprising a extension side and sets a relatively large constant predetermined pressure side damping coefficient side damping coefficient of the shock absorber As a result, the roll motion can be suppressed by obtaining a damping convergence characteristic with respect to the roll input that is equal to or substantially equal to that of the conventional shock absorber, and the feeling of inadequate or excessive feeling of damping force is not given.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a basic configuration of a suspension control device of the present invention.

FIG. 2 is a schematic configuration diagram illustrating an example of a suspension control device of the present invention.

FIG. 3 is a partial cross-sectional front view showing an example of a variable damping force shock absorber employed in the suspension control device of FIG. 2;

FIG. 4 is an enlarged sectional view showing a damping force adjusting mechanism in a maximum damping force state when the vehicle body is lifted.

FIGS. 5A and 5B are enlarged sectional views showing a damping force adjusting mechanism in an intermediate damping force state when the vehicle body is lifted, wherein FIG. 5A shows a hydraulic oil path on the extension side and FIG.

FIGS. 6A and 6B are enlarged cross-sectional views illustrating a damping force adjustment mechanism when the vehicle body does not fluctuate. FIG.

FIGS. 7A and 7B are enlarged cross-sectional views illustrating a damping force adjusting mechanism in a state of a maximum damping force when the vehicle body descends, wherein FIG. 7A illustrates a hydraulic oil path on the extension side and FIG.

FIG. 8 is an explanatory diagram showing a damping force characteristic with respect to a position of a valve body of a variable damping force shock absorber.

FIG. 9 is a block diagram illustrating an example of a controller.

FIG. 10 is an explanatory diagram showing gain characteristics of vibration input and output achieved by a variable damping force shock absorber.

FIG. 11 is an explanatory view for setting a position of a valve body of a basic damping force variable shock absorber by a sprung vertical speed.

FIG. 12 is an explanatory diagram of a damping moment by conventional roll suppression control.

FIG. 13 is an explanatory diagram of a correction coefficient of a control amount for suppressing a bounce movement when a roll movement occurs in the vehicle body posture change suppression control of the embodiment.

FIG. 14 is a flowchart showing one embodiment of a calculation process of vehicle body posture change suppression control executed by the controller of the suspension control device of the present invention.

FIG. 15 is an explanatory diagram of an operation of a bounce motion suppression control which is developed in the vehicle body posture change suppression control by the arithmetic processing of FIG. 14;

16 is an operation explanatory diagram of roll motion suppression control expressed by vehicle body posture change suppression control by the arithmetic processing of FIG. 14;

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1FL-1RR Wheel 2 Body 3FL-3RR Variable damping force variable shock absorber 4 Controller 8 Piston 11 Lower half 12 Upper half 13 Extension oil flow path 14 Pressure oil flow path 31 Valve element 35 Piston rod T1-T3 Extension side flow path C1-C4 Pressure side flow path 41FL-41RR Step motor 51FL-51RR Vertical acceleration sensor 56 Microcomputer 59FL-59RR Motor drive circuit

Claims (1)

(57) [Claims] An actuator disposed between a vehicle body-side member and a wheel-side member and driven in response to an input control signal controls the position of a valve body, thereby allowing the valve to be positioned on one of an extension side and a compression side. A variable damping force shock absorber capable of setting a large damping coefficient or setting both damping coefficients small, a vehicle body motion state detecting means for detecting a motion state caused by a change in posture of the vehicle body, and at least detection by the vehicle body motion state detecting means. A damping coefficient that suppresses a change in the posture of the vehicle body is calculated and set based on the detected vehicle body motion state detection value, and the actual position of the valve body matches the target position of the valve body corresponding to the damping force coefficient. A control unit that outputs the control signal to the actuator to control a damping coefficient of the variable damping force shock absorber. As body movement state detection means, the apparatus further includes a roll movement state detection means for detecting at least a roll movement state occurring in the vehicle, and the control means includes at least a roll movement state detection value detected by the roll movement state detection means. in order to suppress the roll motion state, the damping coefficient of the damping force variable shock absorber which is a extension side and sets a constant predetermined pressure side damping coefficient of the damping coefficient is preset variable damping force shock absorber as the pressure side A suspension control device comprising a roll motion control means for setting a predetermined constant extension side damping coefficient set in advance.